Condensate-free outdoor air cooling unit

ABSTRACT

A highly efficient condensate-free cooling unit functionally based on vapor-compression refrigeration cycle has been described. The condensate collected from the evaporator of the cooling unit is routed through a sub-cooling heat exchanger where it exchanges heat with the primary heat exchange medium emerging through the condenser of the cooling unit, thus, sub-cooling the primary heat exchange medium to a lower temperature before it enters the expansion valve. Emerging from the sub-cooling heat exchanger, the condensate flows through a condensate outlet pipe into multiple spray nozzles disposed over the condensate outlet pipe. The spray nozzles sprinkle the condensate over the hot air blown into the condenser to reduce its temperature. The cooling unit has a substantially higher coefficient of performance compared to the conventional cooling units utilizing vapor-compression refrigeration cycle, and eliminates the problems of condensate removal persistent in the art.

FIELD OF THE INVENTION

The present disclosure deals generally with heat exchange devices, andmore specifically with cooling devices utilizing the vapor compressionrefrigeration cycle.

BACKGROUND OF THE INVENTION

Vapor compression refrigeration cycles are widely used in many coolingsystems for achieving refrigerating effects, including refrigerators,air-conditioning systems, industrial and commercial refrigerationsystems and the like. Typical applications of such cooling units includewarehouses, offices, private and public residential buildings,hospitals, hotels, restaurants and cafeteria. Succinctly, the process ofrefrigeration refers to extracting heat from a space and rejecting itelsewhere, thus lowering the temperature of the space. Vapor compressionrefrigeration cycles use a refrigerant for this process. The cycleincludes four elementary components: a compressor, a condenser, anexpansion valve, and an evaporator. The circulating refrigerant entersthe compressor in the form of saturated vapor and undergoes isentropiccompression, thus increasing its pressure and temperature, andconverting into a superheated vapor. The refrigerant then enters thecondenser, where it comes into thermal contact with cold water or airand rejects heat to it. In the process, the refrigerant absorbs heat(sensible heat) and is converted from superheated vapor to saturatedvapor. After further absorbing latent heat, the saturated vaporeventually is converted to saturated liquid.

At the condenser exit, the refrigerant is thermodynamically a saturatedliquid. That liquid passes through an expansion valve, where it expands,undergoing a reduction in pressure and consequently a partial flashevaporation. This process converts the refrigerant into a mixed liquidand vapor and reduces its temperature to a level below that of the spaceto be refrigerated. The mixture then enters the evaporator where itextracts latent heat from the space and thus completely vaporizes to asaturated vapor. The saturated vapor re-enters the compressor tocomplete the refrigeration cycle. This description pertains to an idealvapor-compression refrigeration cycle, neglecting practical real-worldeffects such as the frictional pressure drop in the system and theslight thermodynamic irreversibility.

In the evaporator, the air to be conditioned passes over heat exchangetubes, exchanging heat with the refrigerant and simultaneously loweringthe air temperature and vaporizing the refrigerant. The ambient airblown over the evaporator's heat exchange tubes generally carries acertain amount of water vapor, and in the course of exchanging heat withthe refrigerant, this vapor partially condenses, forming a watercondensate which starts dripping off from the evaporator.

Appropriate disposal of cooling unit condensate poses a challengingproblem. Providing a separate piping system or condensate collectioncontainers can be burdensome and expensive. Dehumidifiers are employedin certain applications, to reduce the moisture content of the airentering the evaporator. Sometimes, excessive extraction of heat fromthe air even causes frost or ice to build-up in the evaporator and thesurface of its heat exchange tubes, causing the dehumidifier to stop.Thermostats have been used in the art to detect the frost accumulationconditions, by identifying moments when the air temperature drops belowa certain value. The compressor then shuts down until the system warmsup, but those actions shut down the cooling system itself. Efforts havebeen also made in the art for routing the condensate out of theevaporator casing through pumps.

Another challenge is increasing the operating efficiency of the coolingunits. A measure of the operating efficiency of such cycles is thecoefficient of performance (COP), which is the ratio of the obtainableuseful refrigeration effect to the power required to drive the cycle(including the compressor driving power). Either an increase in theuseful refrigeration effect or a decrease in the power input to thecompressor may increase the operating efficiency. Substantiallysuccessful attempts in that direction include disposing an intercoolerbetween the condenser and the expansion valve, routing the highpressure/high temperature refrigerant through the intercooler, where itexchanges heat with the low temperature refrigerant entering theevaporator. This decreases the enthalpy of the high pressure refrigerantand thus increases the coefficient of performance. To increaseefficiency, the added compressor power input must be less than theincrease in the useful refrigeration effect. In some applications, thehigh pressure value of the refrigerant in the cycle is varied and acorresponding change in the value of the COP is observed to derive acorrelation. This correlation is then used to identify the high pressurevalue at which the COP maximizes. Many times, the analysis andidentification of this correlation may be time consuming and inaccurate.

At present, there exists a need for an efficient cooling unit that couldhave a considerably higher coefficient of performance compared to theconventionally used systems, and would simultaneously address theproblem of condensate removal from cooling units, in an effectivemanner.

SUMMARY

The present invention is directed to a highly effective and advantageouscooling unit functionally based on vapor compression refrigerationcycle. The cooling unit has an operable efficiency (coefficient ofperformance) substantially greater than the coefficient of performanceof the conventionally used cooling units, including refrigerators andair-conditioning devices. It further effectively eliminates problem ofcondensate removal persistent in the conventional systems.

In one aspect, a cooling unit for cooling an ambient medium is providedthat includes a compressor, a condenser, an expansion valve, anevaporator, a sub-cooling heat exchanger and a condensate pump. Thecooling unit functionally utilizes vapor compression refrigeration cyclefor achieving refrigerating effect. The evaporator includes a number ofheat exchange tubes through which a primary heat exchange medium flows.The heat exchange tubes remain in continuous thermal communication withthe ambient medium. The condensate pump is connected to the evaporatorto route the condensate collected within the evaporator through thesub-cooling heat exchanger. The sub-cooling heat exchanger has acondensate inlet meant to receive the routed condensate and a condensateoutlet for delivering the condensate to a number of spray nozzlesconnected to it through a condensate outlet pipe. The condensate outletpipe has a set of perforations provided on its surface through which thesub-cooling heat exchanger remains in continuous fluid communicationwith the spray nozzles. Effectively, the condensate is routed from theevaporator to the sub-cooling heat exchanger and further to the spraynozzles. The spray nozzles are directly coupled with hot air blowinginto the condenser and are meant to sprinkle the condensate thereon toreduce its temperature. The sub-cooling heat exchanger further has aprimary heat exchange medium inlet and a primary heat exchange mediumoutlet for a continuous influx and efflux of the primary heat exchangemedium through it. During operations, the condensate and the primaryheat exchange medium remain in continuous thermal communication withinthe sub-cooling heat exchanger and exchange heat with each other.

The cooling unit has a considerably high coefficient of performancecompared to the conventionally used cooling systems. Further, thecooling unit is effectively dry and perfectly addresses the problems ofcondensate removal persistent in the art.

Additional features and advantages of the invention will be madeapparent from the following detailed description of illustrativeembodiments that proceed with reference to the accompanying drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

The summary above, as well as the following detailed description ofpreferred embodiments, is better understood when read in conjunctionwith the appended drawings. For the purpose of illustrating theinvention, exemplary constructions of the invention are shown in thedrawings. The invention is not limited to the specific methods andinstrumentalities disclosed however. Moreover, those in the art willunderstand that the drawings are not to scale. Where possible, likeelements are indicated by identical numbers.

FIG. 1 is the pressure-enthalpy chart for a refrigerant undergoing aconventional vapor compression refrigeration cycle.

FIG. 2 is a schematic diagram showing the basic components of aconventional vapor compression refrigeration system.

FIG. 3 illustrates the different components of an exemplary cooling unitin accordance with the present disclosure

FIG. 4 is another schematic diagram representing the differentcomponents of the cooling unit of FIG. 3

FIG. 5 shows the evaporator of the cooling unit of FIG. 4 in greaterdetails.

FIG. 6 shows the pressure-enthalpy chart for the primary heat exchangemedium circulating within the refrigeration cycle utilized by theexemplary cooling unit of FIG. 4 in accordance with the presentdisclosure.

FIG. 7 illustrates the heat exchange process within the condenser of thecooling unit of FIG. 4.

DETAILED DESCRIPTION OF ILLUSTRATIVE EMBODIMENTS

The description below illustrates embodiments of the claimed invention.This description discloses aspects of the invention but does not defineor limit the invention, such definition and limitation being containedsolely in the claims appended hereto. Those of skill in the art willunderstand that the invention can be implemented in a number of waysdifferent from those set out here, in conjunction with other present orfuture technologies.

As used herein, the following terms carry the had indicated meanings:“Ambient medium” is the medium targeted for conditioning by thedisclosed cooling unit. In a building air-conditioning system, forexample, the ambient medium would be the air inside the building.“Primary heat exchange medium” designates the refrigerant. In mostcircumstances, the primary heat exchange medium would be a refrigerantcapable of exchanging heat with the ambient medium targeted forconditioning.

FIG. 1 shows the pressure-enthalpy diagram for a refrigerant undergoinga typical single-stage vapor compression refrigeration cycle. Thedescription should be read in conjunction with FIG. 2, which depicts thecomponents of a conventional vapor-compression refrigeration system andthe states corresponding to those marked in the pressure-enthalpydiagram for the refrigerant, as shown in FIG. 1. The refrigerant, in theform of saturated vapor at point 1 (pressure P₁ and Temperature T₁),enters the compressor where it is compressed to a high pressure P₂ andhigh temperature T₂, which converts the refrigerant from thethermodynamic state of saturated vapor to superheated vapor at point 2.At this point, the refrigerant exits the compressor. The increase inrefrigerant pressure causes an increase in temperature, thus bringing itto the state 2 where it can be condensed by typically available air orwater.

At state point 2, the refrigerant enters the condenser and flows throughthe condenser coils, where it comes in thermal contact with the air orwater flowing across the coils. The coils allow the refrigerant toreject sensible heat to the air or water, converting it from superheatedvapor to saturated vapor in the course of traversing from point 2 topoint 2. The rejected heat is called sensible heat because its lossleads to a change in refrigerant temperature during the conversion fromsuperheated to saturated vapor. The enthalpy of the refrigerantdecreases in this process as heat is extracted, or equivalently, heat isrejected by it to the air or water in contact. The refrigerant absorbsfurther heat, and loses further enthalpy, until it is converted tosaturated liquid at point 3.

At point 3, the refrigerant exits the condenser and enters the expansionvalve. Here, saturated liquid refrigerant expands, with a sharp loss ofpressure, to point 4 corresponding to the exit from the expansion valve.This sudden decrease in pressure from a value P₃ to P₄ causes acorresponding drop in the temperature of the liquid refrigerant from thecorresponding value T₃ to T₄ and a portion of the liquid refrigerantflashes off into vapor, leaving the refrigerant in mixed liquid andvapor form at the exit 4 of the expansion valve. The enthalpy of therefrigerant however remains constant from point 3 to 4, as shown in FIG.1, as no external heat is added or subtracted.

At point 4, the refrigerant enters the evaporator where it comes intocontact with the air to be cooled. This air flows across the evaporatorcoils and exchanges heat with the refrigerant flowing in the coils. Byextracting the latent heat of vaporization from the air, the refrigerantis converted from a mixed state to saturated vapor at point 1, where itagain enters the compressor and completes the refrigeration cycle.

The coefficient of performance (COP) of a vapor compressionrefrigeration cycle measures the effectiveness of the cycle and isdefined as the ratio of the useful refrigerating effect to the totalpower required to operate the cycle. For the case of a mechanical vaporcompression refrigeration cycle, the total power required is usuallyconsumed in driving the mechanical components including the compressor,the fans, the pumps etc. Mathematically:

$\begin{matrix}{{COP} = \frac{Q_{evaporator}}{W_{net}}} & (1)\end{matrix}$

-   -   where, Q_(evaporator)=net rate of heat extracted by the        evaporator from the ambient medium to be conditioned; and    -   W_(net)=Net power required to drive the cycle components,        including the compressor, the fans and the pumps

As seen from equation (1) above, there are two possible ways to increasethe COP of a vapor compression refrigeration cycle, one by increasingthe rate of energy extraction from the ambient medium to be conditioned(Q_(evap)) and the other by decreasing the power input to the mechanicalcomponents, mainly the compressor (W_(net)).

As shown in FIG. 1, if points 1 and 2 represent the states of therefrigerant during its entry into and exit from the compressor,respectively, and h₁ and h₂ represent the specific enthalpies (enthalpyper unit mass) of the refrigerant at states 1 and 2 respectively, andassuming the process 1→2 being reversible and adiabatic, then the amountof power required to drive the compressor is:

W _(net) =m(h ₂ −h ₁)   (2)

-   -   where m=mass flow rate of the refrigerant at the inlet to the        compressor

The quantity Q_(evap) in eq. (1) depends on the temperature of theambient air flowing into the evaporator, which is uncontrollable to acertain extent, at a specific mass flow rate. Further, the magnitude ofthe higher refrigerant pressure value (P₂) corresponding to point 2 inFIG. 1, depends upon the temperature of the air entering the condenser.The purpose of compressing the refrigerant is to increase its pressure,and hence its temperature, to a value where it can easily reject heat tothe medium entering the condenser. In that sense, the refrigerant has tobe compressed to an extent where the temperature of superheated vaporrefrigerant (T₂) is greater than the temperature of the condenser inletair. If somehow, this extent of compression is reduced (to a pointsomewhere between states 1 and 2), to the point A or point B asrepresented in FIG. 1 for instance, the power required to drive thecompressor would decrease consequently, leading to a tremendous increasein the coefficient of performance of the refrigeration cycle.Specifically, in a case where the refrigerant is compressed to a pointA, instead of point 2, the approximate power input to the compressorwould be:

W ^(/) _(net) =m(h _(a) −h ₁)   (3)

-   -   where h_(a)=specific enthalpy of refrigerant at A and h_(a)<h₂,        W^(/) _(net)<W _(net)

The present disclosure uses this concept to substantially increase thecoefficient of performance of a cooling unit while also effectivelyeliminating the problem of condensate removal.

FIG. 3 is a schematic diagram representing the components of coolingunit 300 in accordance with the present disclosure, as well as the heatexchange process at the evaporator and the condenser side. FIG. 4further shows the different labeled components of the cooling unit shownin FIG. 3. Explaining in conjunction with these figures, cooling unit300 includes an evaporator 410, a compressor 420, a condenser 430, asub-cooling heat exchanger 450, an expansion valve 440, a condensatepump 460 and a number of spray nozzles 480. The primary heat exchangemedium flows along the circuit represented by solid thick lines shown inFIG. 3 and FIG. 4, while condensate flows along the dotted-line circuit.

As illustrated, the primary heat exchange medium (refrigerant), in theform of saturated vapor, enters the compressor 420 at state 1 andconverts into superheated vapor at point 2 ^(/). The superheated vaporrefrigerant rejects heat to air 490 blown into the condenser 430, and asits temperature drops it converts first into saturated vapor and finallyinto saturated liquid at the exit 3 ^(/) from the condenser 430. Therefrigerant is then routed through a sub-cooling heat exchanger 450,where it exchanges heat with condensate 492 collected from theevaporator 410 through the condensate pump 460. Thus, the refrigeranttemperature further decreases and it converts from saturated liquid tosub-cooled liquid at point 3 ^(//). The condensate handling system isdiscussed in detail below. Exiting from the sub-cooling heat exchanger,the refrigerant enters expansion valve 440, where it expands andundergoes an abrupt decrease in pressure, thus lowering its temperature.A fraction of the liquid refrigerant undergoes partial flashevaporation, and it exits expansion valve 440 as a mixed liquid andvapor at point 4 ^(/). Entering the evaporator 410, the refrigerant, itstemperature substantially lowered, extracts heat from ambient air 495,and vaporizes completely before exiting the evaporator 410 at state 1.The vaporized refrigerant reenters the compressor 420 to complete therefrigeration cycle.

The ambient air 495 generally has a certain absolute humidity, and theheat extraction process generally results in at least a portion of thatmoisture condensing as condensate 492 within the evaporator 410.Condensate pump 460 routes this condensate 492 along a condensate flowpath K→L→M as represented by the dotted lines in FIG. 4. Following thispath, the condensate 492 enters the sub-cooling heat exchanger 450through a condensate inlet 452, exchanges heat with the refrigerantflowing through the sub-cooling exchanger 450, and finally exits througha condensate outlet 454. Thereafter, the condensate 492 flows outthrough a condensate outlet pipe 485 connected to the condensate outlet454. A number of spray nozzles 480 are mounted spatially equidistant onthe condensate outlet pipe 485, and set of perforations 486 conveycondensate 492 to the spray nozzles 480.

The spray nozzles 480 are arranged to continuously spray condensate 492into the air stream 490 flowing into condenser 430. That flow can beassisted by a blower fan 470, and the condensate spray acts on airstream 490 to reduce its temperature, as is clear to those of skill inthe art. As a result, air stream 490 makes contact with the condensercoils at a reduced temperature, enhancing its ability to extract heatfrom the refrigerant.

Compressor 420 could be any suitable compressor known in the art usablefor refrigerating units, and the selection of the exact compressor typeis based on the operating conditions of the cooling unit 300 and certainassociated design and maintenance criteria. To minimize losses in therefrigerant pressure and reduce the maintenance activities, and forcases where prolonged maintenance-free operations of the cooling unit300 is desirable, a hermetically sealed or a semi-hermetic compressorwould be preferred. An electric motor (not shown) is used a source ofcontinuous power supply to the compressor 420 during operatingconditions. Further, compressor 420 can be either a positivedisplacement or dynamic compressor depending upon the cooling loadrequired and the environment of installation of the cooling unit 300. Inlarge buildings and warehouses demanding huge cooling capacity, adynamic type single-stage centrifugal compression would be preferred.Cases wherein extremely high output pressures at the compressor outletare required, a multi-staged centrifugal compressor would be preferablyused. In certain embodiments, a variable displacement type compressormay be used wherein the mass flow rate of the refrigerant needs to befrequently controlled based on variations in the temperature of the airblown into the evaporator 410.

Any suitable refrigerant known to those in the art can be used as theprimary heat exchange medium for circulation within the cooling unit 300for practicing the present disclosure. Common examples include thewell-known family of refrigerants denoted by the ‘R-number’ systemincluding R-11, R-22, R-134 (a) and others.

The sub-cooling heat exchanger 450 is be designed to provide sufficientcapacity to incorporate high volume of condensate 492 generated duringprolonged operations of the cooling unit 300. Those skilled in the artwould recognize that the size and capacity of the sub-cooling heatexchanger 450 can be varied based on the desired cooling capacity of thecooling unit 300, as well as upon the expected conditions of ambientair. During inoperative conditions of the cooling unit 300, thesub-cooling heat exchanger 450 would act as a reservoir for thecondensate 492 collected during former operations. Further, the apertureof the condensate outlet 454 is preferably smaller than the aperture ofthe condensate inlet 452, so that the volume of condensate entering thesub-cooling heat exchanger 450 is more than the volume leaving it. Inthat manner, a certain volume of condensate accumulates in thesub-cooling heat exchanger 450, and thus the level of condensate risesduring continuing operations of the cooling unit 300. The condensateinlet 452 is provided at a higher elevation with respect to thecondensate outlet 454 to ensure that there is a continuous influx of thecondensate 492 within the sub-cooling heat exchanger 450. Eventually,with the elevated level of the accumulated condensate 492 in thesub-cooling heat exchanger 450, the condensate 492 flows out through thecondensate outlet 454 into the condensate outlet pipe 485 by virtue ofthe achieved velocity of efflux. Further, the pressurized influx of thecondensate 492 into the sub-cooling heat exchanger 450 by the condensatepump 460 provides additional expelling impulse for the volume ofcondensate 492 leaving the sub-cooling heat exchanger through thecondensate outlet 454.

Condensate pump 460 pressurizes the condensate 492 to flow through thesub-cooling heat exchanger 450 and on to spray nozzles 480. Thecondensate velocity in that flow circuit should be sufficient to allowthe spray nozzles 480 to discharge the condensate as a spray of finedroplets, maximizing heat exchange between the condensate and the airstream 490. To minimize energy consumption, hydraulic spray nozzles arepreferred. However, those skilled in the art would understand that anyother type of spray nozzles conventionally known in the art may also beused, including gas atomized spray nozzles, thus, not limiting the scopeof the invention.

The sub-cooling heat exchanger 450 further includes a primary heatexchange medium inlet 455 and a primary heat exchange medium outlet 456for a continuous influx and discharge of refrigerant. Exiting thecondenser 430, the refrigerant enters the sub-cooling heat exchanger 450inlet 455, flows through a set of cooling coils 458, and finally emergesthrough outlet 456. That flow allows the refrigerant to reject heat tothe condensate 492, and to emerge at point 3 ^(//) as a sub-cooledliquid. This decreases the temperature of the refrigerant as it exitsthe sub-cooling heat exchanger 450. This heat exchange is an isobaricprocess, occurring at the constant saturation pressure of the saturatedliquid refrigerant exiting the condenser 430 at point 3 ^(/). Thespecific enthalpy of the refrigerant decreases as it moves from thestate 3 ^(/) to the state 3 ^(//).

FIG. 5 illustrates the evaporator 410 to be used in the cooling unit 300of FIG. 4, in further detail. Evaporator 410 includes a casing 501having an anterior surface 502 and a posterior surface 503. A set ofheat exchange tubes 504 are disposed in parallel within the casing 501.The heat exchange tubes 504 extend along the entire length of the casing501, lying substantially perpendicular to the anterior and posteriorsurfaces 502 and 503 respectively. Further, the anterior surface 502 andposterior surface 503 have perforations through which the heat exchangetubes extend into and out of casing 501. Fins 505 are equidistantlydisposed within the casing over the heat exchange tubes 504. Thoseskilled in the art would recognize that different evaporator structurescould easily be substituted for the disclosed component without alteringthe performance of the disclosed embodiment.

FIG. 6 illustrates the pressure-enthalpy diagram for the refrigerantcirculating in the cooling unit 300 of the present disclosure. The solidlined path 1→2 ^(/)→2 ^(//)→3 ^(/)→3 ^(//)→4 ^(/)>1 represents therespective states marked in FIG. 3 and FIG. 4, for the refrigerantflowing in the cooling unit 300 according to the present disclosure.Path 1→2 ^(/)→2→3→4→1 (dotted lines) represents the pressure-enthalpyvariation of the refrigerant if it flows through a conventional coolingunit based on the vapor-compression refrigeration cycle and not equippedwith a sub-cooling heat exchanger and spray nozzles, as shown by FIG. 2.The horizontal axis OX represents the specific enthalpy (enthalpy perunit mass) of the refrigerant at different thermodynamic state points,while the vertical axis OB represents the pressure of the refrigerant atdifferent states. The two superimposed paths lay out a comparisonbetween conventional cooling units and the inventive cooling unit 300 ofFIG. 4.

As shown, in conventional refrigerating units, the air at the inlet tothe condenser is at a relatively higher temperature, and hence therefrigerant is compressed to a higher pressure P₂ and correspondingly ahigher temperature T₂, so that it can easily reject heat to the airentering the condenser. In the inventive cooling unit 300, the air blowninto the condenser 430 is sprinkled with condensate 492 by the spraynozzles 480 (FIG. 4), before that air stream enters the condenser 430.Hence, the temperature of the air 490 entering the condenser 430 isrelatively lower. Therefore, the refrigerant in the cooling unit 300needs to be compressed to a lower pressure represented by a point 2 ^(/)lying somewhere on the curve joining the points 1 and 2 in FIG. 6. Theexact location of this point 2 ^(/) on the curved line 1→2 depends oncertain factors, the primary ones being the temperature of the air 490entering the condenser 430, its mass flow rate and the mass flow rate ofthe refrigerant circulating in the cooling unit 300.

Further, during process 3 ^(/)→3 ^(//), the refrigerant flows throughthe sub-cooling heat exchanger where it rejects heat to the condensateand emerges as a sub-cooled liquid at point 3 ^(//). Process 3 ^(//)→4^(/) represents the expansion of the refrigerant within the expansionvalve 440. This expansion abruptly decreases the pressure of therefrigerant from P₂ ^(/) to P₁, though the enthalpy of the refrigerantremains at a constant value h₄ ^(/) during the process. In comparison,the refrigerant in conventional cooling unit of FIG. 2 would undergo thepath 2→3 in the condenser, emerge as a saturated liquid at point 3, andfinally expand in the expansion valve isobarically to the point 4(process 3→4 in FIG. 6), consequently decreasing its pressure fromP_(2 to P) _(1.)

Portion 510 on axis OX in FIG. 6 represents the decrease in the requiredpower input to the compressor achieved by the inventive cooling unit 300of FIG. 4, in comparison to a conventional cooling unit of FIG. 2. Asset out earlier in Eq. (2), the power required to drive the compressorof a conventional cooling unit of FIG. 2 is:

W _(net) ={dot over (m)}(h ₂ −h ₁)   (2)

-   -   where, {dot over (m)}=mass flow rate of the refrigerant flowing        in the vapor-compression refrigeration cycle.

Using the same refrigerant, at the same mass flow rate {dot over (m)},the power required to drive the compressor of the inventive cooling unitof FIG. 4 is:

W ^(/) _(net) ={dot over (m)}(h ₂ ^(/) −h ₁)   (3)

Since h₂ ^(/)<h₂ (as seen in FIG. 5), W^(/) _(net)<W_(net)

The decreased power input to the compressor, as achieved by the coolingunit 300 of FIG. 4 is obtainable by subtracting Eq. (3) from Eq. (2),and is:

ΔW=W _(net) −W ^(/) _(net) ={dot over (m)}(h ₂ −h ₂ ^(/))   (4)

The length of the portion 510 on the enthalpy axis OX of FIG. 5 actuallyrepresents this decreased compressor power input per unit mass flow rateof the refrigerant (ΔW/{dot over (m)}), as achieved by the cooling unit300 of FIG. 4.

Further, portion 520 in FIG. 6 represents the enhanced cooling effectachieved by the cooling unit 300 of FIG. 4. This enhanced cooling effect520 is equivalent to the additional heat energy extracted by theevaporator 410 of cooling unit 300 (denoted by Q^(/) _(evap)) from theambient air 495 to be conditioned, in comparison to the heat actuallyextracted by the evaporator of a conventional cooling unit of FIG. 2(denoted by Q_(evap)).

For a conventional cooling unit of FIG. 2, the refrigerant undergoes theprocess 4→1 (shown in FIG. 6) in the evaporator and hence, the heatgained by the refrigerant from the ambient air to be conditioned is:

Q _(evap) ={dot over (m)}(h ₁ −h ₄)   (5)

In the inventive cooling unit 300, the refrigerant is further sub-cooledto state 3 ^(//) and hence undergoes the process represented by 4 ^(/)→1within the evaporator. Therefore, the heat gained by the refrigerant is:

Q ^(/) _(evap) ={dot over (m)}(h ₁ −h ₄ ^(/))   (6)

Since, h₄ ^(/)<h₄, we have Q^(/) _(evap)>Q_(evap)

The approximate enhanced cooling effect gained by the cooling unit 300,can be obtained from Eq. (5) and Eq. (6) above, and is:

ΔQ _(evap) =Q ^(/) _(evap) −Q _(evap) ={dot over (m)}(h ₄ −h ₄ ^(/))  (7)

The length of the portion 520 on the enthalpy axis OX of FIG. 5 actuallyrepresents this enhanced cooling effect per unit mass flow rate of therefrigerant, i.e., ΔQ_(evap)/{dot over (m)}.

The coefficient of performance of the cooling unit 300 of FIG. 4 issubstantially increased in this manner, compared with the COP of theconventional cooling units utilizing vapor-compression refrigerationcycles. As stated in eq. (i) earlier, the COP of a refrigeration cycleis given by:

$\begin{matrix}{{COP} = \frac{Q_{evaporator}}{W_{net}}} & (1)\end{matrix}$

Referring to FIG. 6, the value of COP for a conventional unit of FIG. 3would be:

$\begin{matrix}\begin{matrix}{{COP} = \frac{Q_{evaporator}}{W_{net}}} \\{= \frac{\left( {h_{1} - h_{4}} \right)}{\left( {h_{2} - h_{1}} \right)}}\end{matrix} & {(8)\mspace{14mu}\left\lbrack {{from}\mspace{14mu} {eq}\mspace{14mu} (2)\mspace{14mu} {and}\mspace{14mu} (5)} \right\rbrack}\end{matrix}$

For the inventive cooling unit 300 of FIG. 4, the enhanced COP is givenby:

$\begin{matrix}{{COP}^{/} = \frac{\left( {h_{1} - h_{4}^{/}} \right)}{\left( {h_{2}^{/} - h_{1}} \right)}} & {(9)\mspace{14mu}\left\lbrack {{from}\mspace{14mu} {eq}\mspace{14mu} (3)\mspace{14mu} {and}\mspace{14mu} (4)} \right\rbrack}\end{matrix}$

Equations (8) and (9) clearly depict the increase in Coefficient ofperformance obtained by the cooling unit 300 of FIG. 4. For the samerefrigerant and the same value its mass flow rate in the refrigerationcycle, the cooling unit 300 obtains a substantially enhanced coolingcapacity with a considerably lower value of the power required to drivethe compressor 420, as compared to the conventional cooling unit of FIG.3. Additionally, the problem of condensate removal from the coolingunits based on vapor-compression refrigeration cycles is significantlyalleviated by the inventive cooling unit 300.

FIG. 7 further illustrates the heat transfer phenomenon within thecondenser of the cooling unit 300 of FIG. 4 in accordance with thepresent disclosure. As shown, control volume 700 surrounds the condenser430 and all the heat exchange processes at the condenser side occurwithin the control volume 700. Since the condenser 430 is air-cooled,hot outdoor air 710 acquires an average temperature T₁ after thecondensate 492 is sprinkled into the airstream, enters the controlvolume 700 and leaves the condenser 430 at a higher average temperatureT_(m) after gaining heat from the refrigerant. Refrigerant 730 entersthe condenser 430 at a pressure P₂ ^(/) and temperature T₂ ^(/)(corresponding to the state 2 ^(/) marked in FIG. 6) and emerges at thesame pressure but a lower temperature T₃ ^(/) after rejecting heat tothe hot air 710 blown into the condenser 430. The temperature values T₁and T_(m) for the hot air entering and exiting the control volume 700are measurable through any suitable temperature detecting device knownin the art, and conventionally used for the purpose, including athermocouple, a thermostat etc. Neglecting the heat gained by theinteriors of the condenser 430, including the coil tubes through whichthe refrigerant 730 flows, and the minute heat losses to thesurroundings, the major part of the heat lost by the refrigerant 730goes into the hot air 710 entering the condenser 430, thus resulting inan elevation in its temperature from an average inlet temperature valueT₁ to the outlet value T_(m).

-   -   Let {dot over (m)}_(air)=mass flow rate of the air blown into        the condenser;    -   {dot over (m)}_(ref)=mass flow rate of the refrigerant through        the vapor-compression cycle for the cooling unit 300    -   C=average specific heat capacity of the hot air within        temperature range T₁ to T_(m)    -   h₂ ^(/), h₃ ^(/)=enthalpies of the refrigerant at entry point 2        ^(/) and exit point 3 ^(/), as marked in the pressure-enthalpy        diagram of FIG. 6    -   q_(ref)=average rate of heat lost by the refrigerant 730 within        the control volume 700    -   q_(air)=average rate of heat gained by the hot air 710 within        the control volume 700    -   We have the following approximations under above mentioned        considerations:

q _(ref) ={dot over (m)} _(ref)(h ₂ ^(/) −h ₃ ^(/))   (10)

q _(air) ={dot over (m)} _(air) C(T _(m) −T ₁)   (11)

-   -   Neglecting the minor losses within the condenser:

q_(ref)=q_(air)

{dot over (m)} _(ref)(h ₂ ^(/) −h ₃ ^(/))={dot over (m)} _(air) C(T _(m)−T ₁) [By equating eqns. (10) and (11)]

(h ₂ ^(/) −h ₃ ^(/))={dot over (m)} _(air) C(T _(m) −T ₁)/{dot over (m)}_(ref)   (12)

This calculated value of the change in enthalpy of the refrigerantduring its flow through the condenser is used as a reference forapproximating the extent of the point 2 ^(/) (shown in FIG. 6) untilwhich the refrigerant needs to be compressed by the compressor 420 ofthe cooling unit 300.

For a specific temperature (T₁) of the air 710 at the inlet to thecondenser 700, the value of the higher pressure P₂ ^(/) to which therefrigerant should undergo compression in the compressor 420, formaximizing the cooling capacity of the cooling unit, is obtainable bythe standard pressure-enthalpy chart for the refrigerant used. Withknown values of the mass flow rate of the refrigerant ({dot over(m)}_(ref)) and the air ({dot over (m)}_(air)), and an approximate valueof the specific heat capacity of the air in the temperature range T₁ toT_(m), the corresponding difference between the enthalpies of thesuperheated vapor refrigerant (entering the condenser) and the saturatedliquid refrigerant (exiting the condenser), i.e., (h₂ ^(/)−h₃ ^(/)), iscalculated from Eq. (12).

The high pressure value of the refrigerant (P₂ ^(/)) is calibrated to bein conformity with the enthalpy difference calculated in eq. (xii) usingthe standard pressure-enthalpy chart for the refrigerant used, and thiscalibrated value P₂ ^(/) is set for the compressor 420 of the coolingunit 300. As an example, using R-134 (a) as the working refrigerantfluid, with the high pressure side saturated vapor temperature of 50° C.(considering hot summer outdoor air entering the condenser), theenthalpies of the saturated liquid refrigerant and the saturated vaporrefrigerant are found to be 270 KJ/Kg. K and 420 KJ/Kg. K respectively,from the standard pressure-enthalpy chart for R-134 (a). Consideringspecific operating values of the mass flow rate of the hot air and therefrigerant in Eq. (12), if the enthalpy difference h₂ ^(/)−hd 3 ^(/) is200 KJ/Kg K (for instance), then the refrigerant needs to be compressedto a pressure of about 1.4 Mpa (P₂ ^(/)), and to a temperature of about90° C. (T₂ ^(/)), for proper functioning of the cooling unit 300.

Although the present invention has been described in considerabledetails with reference to certain preferred versions thereof, otherversions are also possible.

The cooling unit as disclosed herein can be used in severalcircumstances where a refrigerating effect is desired. In an aspect, thecooling unit can be an integral part of a usual air-conditioning systemsutilized in homes or other buildings. As another example, several suchcooling units can be simultaneously used in collaboration for commercialand industrial applications where large scale air-conditioning isrequired, including residential buildings, factories etc. As a furtherexample, the unit can also be used in conditioning the air in movietheatres, concert halls, restaurants, cafeteria etc. The appropriatemethod of use would be to install the evaporator of the cooling unit ata suitable location within the space where the refrigerating effect isdesired such that it can extract heat from the space, condition the airand reject this heat elsewhere. These and other variations are wellwithin the scope of those of ordinary skill in the art.

1. A cooling unit employing vapor compression refrigeration cycle forcooling an ambient medium, the cooling unit comprising: an evaporator,including a plurality of heat exchange tubes adapted to carry a primaryheat exchange medium, and arranged to receive a flow of an ambientmedium; a condensate pump positioned to route condensate from theevaporator; a sub-cooling heat exchanger in fluid communication with thecondensate pump; and a plurality of spray nozzles in simultaneous fluidcommunication with the sub-cooling heat exchanger and positioned tospray condensate into the flow of the ambient medium.
 2. The coolingunit of claim 1, wherein the sub-cooling heat exchanger has a condensateinlet for a continuous inflow of the condensate therein, and acondensate outlet for a continuous outflow of the condensate therefrom.3. The cooling unit of claim 2, wherein the condensate outlet has across-sectional area preferably smaller than the cross-sectional area ofthe condensate inlet.
 4. The cooling unit of claim 2, wherein thecondensate inlet is positioned at a higher elevation with respect to thecondensate outlet.
 5. The cooling unit of claim 2, wherein thesub-cooling heat exchanger is in fluid communication with a condensateoutlet pipe, the condensate outlet pipe having the spray nozzlesmounted.thereon.
 6. The cooling unit of claim 5, wherein the condensateoutlet pipe is provided with a set of perforations to fluidlycommunicate with the plurality of spray nozzles.
 7. The cooling unit ofclaim 1, wherein the sub-cooling heat exchanger has a primary heatexchange medium inlet for a continuous inflow of the primary heatexchange medium therein, and a primary heat exchange medium outlet for acontinuous outflow of the primary heat exchange medium therefrom.
 8. Thecooling unit of claim 1, wherein the primary heat exchange medium andthe condensate are in thermal communication within the sub-cooling heatexchanger.
 9. The cooling unit of claim 1, wherein the sub-cooling heatexchanger is in fluid communication with least one of the plurality ofspray nozzles.
 10. The cooling unit of claim 1, wherein the plurality ofspray nozzles are arranged in a spaced array.
 11. The cooling unit ofclaim 1, wherein the plurality of spray nozzles are coupled to acondenser, and configured to continuously spray the condensate on airblown into the condenser.
 12. A method of increasing the coefficient ofperformance of a cooling unit utilizing vapor compression refrigerationcycle, the cooling unit including an evaporator, a compressor, acondenser and an expansion valve, the method comprising: collectingcondensate from the evaporator; routing condensate to a sub-cooling heatexchanger; exchanging heat in the sub-cooling heat exchanger between thecondensate and a primary heat exchange medium; directing the condensatefrom the sub-cooling heat exchanger to a plurality of spray nozzles; andsprinkling the condensate through the plurality of spray nozzles overair blown into the condenser.
 13. The method of claim 12 wherein thecooling unit includes a compressor, and a condenser configured toreceive a flow of an ambient medium, the method further comprisingcompressing the primary heat exchange medium within the compressor to ahigh-pressure value based at least on a set of parameters correspondingto the ambient medium.
 14. The method of claim 13, wherein the set ofparameters includes a mass-flow rate value of the ambient medium flowingthrough the condenser.
 15. The method of claim 13, wherein the set ofparameters include the corresponding average temperature values of theambient medium entering and leaving the condenser.
 16. The method ofclaim 12, wherein the condensate remains in continuous thermalcommunication with a primary heat exchange medium within the sub-coolingheat exchanger.
 17. The method of claim 12 further comprising directingthe condensate through a condensate outlet pipe from the sub-coolingheat exchanger to the spray nozzles.
 18. The method of claim 17 furthercomprising providing a set of perforations on the condensate outlet pipeto enable it being in simultaneous thermal communication with theplurality of spray nozzles.